反鏟式單斗液壓挖掘機工作裝置設計及其運動分析設計

喜歡這套資料就充值下載吧。資源目錄里展示的都可在線預覽哦。下載后都有，請放心下載，文件全都包含在內(nèi)，有疑問咨詢QQ:414951605 或 1304139763 目 錄前言 1一、緒論2（一）國內(nèi)外研究狀況2（二）論文構(gòu)成及研究內(nèi)容2二、總體方案設計3（一）工作裝置構(gòu)成3（二）動臂及斗桿的結(jié)構(gòu)形式5（三）動臂油缸與鏟斗油缸的布置5（四）鏟斗與鏟斗油缸的連接方式5（五）鏟斗的結(jié)構(gòu)選擇6（六）原始幾何參數(shù)的確定7三、工作裝置運動學分析8（一）動臂運動分析8（二）斗桿的運動分析10（三）鏟斗的運動分析11（四）特殊工作位置計算15四、挖掘阻力分析18（一）轉(zhuǎn)斗挖掘阻力計算18（二）斗桿挖掘阻力計算18五、基本尺寸的確定 20（一）斗形參數(shù)的確定20（二）動臂機構(gòu)參數(shù)的選擇201、 1與A點坐標的選取202、 l1與l2的選擇203、 l41與l42的計算214、 l5的計算 21（三）動臂機構(gòu)基本參數(shù)的校核 231、 動臂機構(gòu)閉鎖力的校核232、 滿斗處于最大挖掘半徑時動臂油缸提升力矩的校核253、 滿斗處于最大高度時,動臂提升力矩的校核 26（四）斗桿機構(gòu)基本參數(shù)的選擇27（五）鏟斗機構(gòu)基本參數(shù)的選擇281、 轉(zhuǎn)角范圍282、 鏟斗機構(gòu)其它基本參數(shù)的計算28六、工作裝置結(jié)構(gòu)設計 30（一）斗桿的結(jié)構(gòu)設計301、斗桿的受力分析 302、斗桿內(nèi)力圖的繪制 353、 結(jié)構(gòu)尺寸的計算37（二）動臂結(jié)構(gòu)設計391、危險工況受力分析 422、內(nèi)力圖和彎矩圖的求解 433、 結(jié)構(gòu)尺寸的計算45（三）鏟斗的設計471、鏟斗斗形尺寸的設計 472、鏟斗斗齒的結(jié)構(gòu)計算 473、 鏟斗的繪制48七、銷軸與襯套的設計 49（一）銷軸的設計49（二）銷軸用螺栓的設計49（三）襯套的設計49八、總結(jié)50九、參考文獻 51十、致謝52附件一 外文翻譯 53目 錄前言 1一、緒論2（一）國內(nèi)外研究狀況2（二）論文構(gòu)成及研究內(nèi)容2二、總體方案設計3（一）工作裝置構(gòu)成3（二）動臂及斗桿的結(jié)構(gòu)形式5（三）動臂油缸與鏟斗油缸的布置5（四）鏟斗與鏟斗油缸的連接方式5（五）鏟斗的結(jié)構(gòu)選擇6（六）原始幾何參數(shù)的確定7三、工作裝置運動學分析8（一）動臂運動分析8（二）斗桿的運動分析10（三）鏟斗的運動分析11（四）特殊工作位置計算15四、挖掘阻力分析18（一）轉(zhuǎn)斗挖掘阻力計算18（二）斗桿挖掘阻力計算18五、基本尺寸的確定 20（一）斗形參數(shù)的確定20（二）動臂機構(gòu)參數(shù)的選擇201、 1與A點坐標的選取202、 l1與l2的選擇203、 l41與l42的計算214、 l5的計算 21（三）動臂機構(gòu)基本參數(shù)的校核 231、 動臂機構(gòu)閉鎖力的校核232、 滿斗處于最大挖掘半徑時動臂油缸提升力矩的校核253、 滿斗處于最大高度時,動臂提升力矩的校核 26（四）斗桿機構(gòu)基本參數(shù)的選擇27（五）鏟斗機構(gòu)基本參數(shù)的選擇281、 轉(zhuǎn)角范圍282、 鏟斗機構(gòu)其它基本參數(shù)的計算28六、工作裝置結(jié)構(gòu)設計 30（一）斗桿的結(jié)構(gòu)設計301、斗桿的受力分析 302、斗桿內(nèi)力圖的繪制 353、 結(jié)構(gòu)尺寸的計算37（二）動臂結(jié)構(gòu)設計391、危險工況受力分析 422、內(nèi)力圖和彎矩圖的求解 433、 結(jié)構(gòu)尺寸的計算45（三）鏟斗的設計471、鏟斗斗形尺寸的設計 472、鏟斗斗齒的結(jié)構(gòu)計算 473、 鏟斗的繪制48七、銷軸與襯套的設計 49（一）銷軸的設計49（二）銷軸用螺栓的設計49（三）襯套的設計49八、總結(jié)50九、參考文獻 51十、致謝52附件一 外文翻譯 53反鏟式單斗液壓挖掘機工作裝置設計及其運動分析設計引 言挖掘機在國民經(jīng)濟建設的許多行業(yè)被廣泛地采用，如工業(yè)與民用建筑、交通運輸、水利電氣工程、農(nóng)田改造、礦山采掘以及現(xiàn)代化軍事工程等等行業(yè)的機械化施工中。據(jù)統(tǒng)計，一般工程施工中約有60%的土方量、露天礦山中80%的剝離量和采掘量是用挖掘機完成的。隨著我國基礎設施建設的深入和在建設中挖掘機的廣泛應用，挖掘機市場有著廣闊的發(fā)展空間，因此發(fā)展?jié)M足我國國情所需要的挖掘機是十分必要的。而工作裝置作為挖掘機的重要組成部分，對其研究和控制是對整機開發(fā)的基礎。反鏟式單斗液壓挖掘機工作裝置是一個較復雜的空間機構(gòu)，國內(nèi)外對其運動分析、機構(gòu)和結(jié)構(gòu)參數(shù)優(yōu)化設計方面都作了較深入的研究，具體的設計特別是中型挖掘機的設計已經(jīng)趨于成熟。關于反鏟式單斗液壓挖掘機的相關文獻也很多，這些文獻從不同側(cè)面對工作裝置的設計進行了論述。而筆者的設計知識和水平還只是一個學步的孩子，進行本課題的設計是為對挖掘機的工作裝置設計有一些大體的認識，掌握實際工程設計的流程、方法，鞏固所學的知識和提高設計能力。一、緒論(一）國內(nèi)外研究狀況當前，國際上挖掘機的生產(chǎn)正向大型化、微型化、多能化和專用化的方向發(fā)展。國外挖掘機行業(yè)重視采用新技術、新工藝、新結(jié)構(gòu)和新材料，加快了向標準化、系列化、通用化發(fā)展的步伐。我國己經(jīng)形成了挖掘機的系列化生產(chǎn)，近年來還開發(fā)了許多新產(chǎn)品，引進了國外的一些先進的生產(chǎn)率較高的挖掘機型號。由于使用性能、技術指標和經(jīng)濟指標上的優(yōu)越，世界上許多國家，特別是工業(yè)發(fā)達國家，都在大力發(fā)展單斗液壓挖掘機。目前，單斗液壓挖掘機的發(fā)展著眼于動力和傳動系統(tǒng)的改進以達到高效節(jié)能;應用范圍不斷擴大，成本不斷降低，向標準化、模塊化發(fā)展，以提高零部件、配件的可靠性，從而保證整機的可靠性;電子計算機監(jiān)測與控制，實現(xiàn)機電一體化;提高機械作業(yè)性能，降低噪音，減少停機維修時間，提高適應能力，消除公害，縱觀未來，單斗液壓挖掘機有以下的趨勢:1、向大型化發(fā)展的同時向微型化發(fā)展。2、更為普遍地采用節(jié)能技術。3、不斷提高可靠性和使用壽命。4、工作裝置結(jié)構(gòu)不斷改進，工作范圍不斷擴大。5、由內(nèi)燃機驅(qū)動向電力驅(qū)動發(fā)展。6、液壓系統(tǒng)不斷改進，液壓元件不斷更新。7、應用微電子、氣、液等機電一體化綜合技術。8、增大鏟斗容量，加大功率，提高生產(chǎn)效率。9、人機工程學在設計中的充分利用。（二）論文構(gòu)成及研究內(nèi)容本論文主要對由動臂、斗桿、鏟斗、銷軸、連桿機構(gòu)組成挖掘機工作裝置進行設計。具體內(nèi)容包括以下五部分:1、 挖機工作裝置的總體設計。2、 挖掘機的工作裝置詳細的機構(gòu)運動學分析。3、 工作裝置各部分的基本尺寸的計算和驗證。4、 工作裝置主要部件的結(jié)構(gòu)設計。5、 銷軸的設計及螺栓等標準件進行選型。二、總體方案設計（一）工作裝置構(gòu)成1-斗桿油缸；2- 動臂； 3-油管； 4-動臂油缸； 5-鏟斗； 6-斗齒； 7-側(cè)板；8-連桿； 9-曲柄： 10-鏟斗油缸； 11-斗桿圖2.1 工作裝置組成圖 圖2.1為液壓挖掘機工作裝置基本組成及傳動示意圖，如圖所示反鏟工作裝置由鏟斗5、連桿9、斗桿11、動臂2、相應的三組液壓缸1, 4,10等組成。動臂下鉸點鉸接在轉(zhuǎn)臺上，通過動臂缸的伸縮，使動臂連同整個工作裝置繞動臂下鉸點轉(zhuǎn)動。依靠斗桿缸使斗桿繞動臂的上鉸點轉(zhuǎn)動，而鏟斗鉸接于斗桿前端，通過鏟斗缸和連桿則使鏟斗繞斗桿前鉸點轉(zhuǎn)動。挖掘作業(yè)時，接通回轉(zhuǎn)馬達、轉(zhuǎn)動轉(zhuǎn)臺，使工作裝置轉(zhuǎn)到挖掘位置，同時操縱動臂缸小腔進油使液壓缸回縮，動臂下降至鏟斗觸地后再操縱斗桿缸或鏟斗缸，液壓缸大腔進油而伸長，使鏟斗進行挖掘和裝載工作。鏟斗裝滿后，鏟斗缸和斗桿缸停動并操縱動臂缸大腔進油，使動臂抬起，隨即接通回轉(zhuǎn)馬達，使工作裝置轉(zhuǎn)到卸載位置，再操縱鏟斗缸或斗桿缸回縮，使鏟斗翻轉(zhuǎn)進行卸土。卸完后，工作裝置再轉(zhuǎn)至挖掘位置進行第二次挖掘循環(huán)。在實際挖掘作業(yè)中，由于土質(zhì)情況、挖掘面條件以及挖掘機液壓系統(tǒng)的不同，反鏟裝置三種液壓缸在挖掘循環(huán)中的動作配合可以是多樣的、隨機的。上述過程僅為一般的理想過程。挖掘機工作裝置的大臂與斗桿是變截面的箱梁結(jié)構(gòu)，鏟斗是由厚度薄的鋼板焊接而成。各油缸可看作是只承受拉壓載荷的桿。根據(jù)以上特征，可以對工作裝置進行適當簡化處理。則可知單斗液壓挖掘機的工作裝置可以看成是由動臂、斗桿、鏟斗、動臂油缸、斗桿油缸、鏟斗油缸及連桿機構(gòu)組成的具有三自由度的六桿機構(gòu)，處理的具體簡圖如2.2所示。進一步簡化得圖如2.3所示。圖2.2 工作裝置結(jié)構(gòu)簡圖1-鏟斗；2-連桿；3-斗桿；4-動臂；5-鏟斗油缸；6-斗桿油缸圖2.3 工作裝置結(jié)構(gòu)簡化圖挖掘機的工作裝置經(jīng)上面的簡化后實質(zhì)是一組平面連桿機構(gòu)，自由度是3，即工作裝置的幾何位置由動臂油缸長度L1、斗桿油缸長度L2、鏟斗油缸長度L3決定，當L1、L2、L3為某一確定的值時，工作裝置的位置也就能夠確定。(二）動臂及斗桿的結(jié)構(gòu)形式動臂采用整體式彎動臂，這種結(jié)構(gòu)形式在小型挖掘機中應用較為廣泛。其結(jié)構(gòu)簡單、價廉，剛度相同時結(jié)構(gòu)重量較組合式動臂輕，且有利于得到較大的挖掘深度。斗桿也有整體式和組合式兩種，大多數(shù)挖掘機采用整體式斗桿。在本設計中由于不需要調(diào)節(jié)斗桿的長度，故也采用整體式斗桿。（三）動臂油缸與鏟斗油缸的布置動臂油缸裝在動臂的前下方，動臂的下支承點（即動臂與轉(zhuǎn)臺的鉸點）設在轉(zhuǎn)臺回轉(zhuǎn)中心之前并稍高于轉(zhuǎn)臺平面，這樣的布置有利于反鏟的挖掘深度。大部分中小型液壓挖掘機以反鏟作業(yè)為主，常采用動臂支點靠前布置的方案。油缸活塞桿端部與動臂的鉸點設在動臂箱體下底板的凸緣上，雖然這樣會影響動臂的下降幅度，但不會削弱動臂的結(jié)構(gòu)強度，而且使動臂的受力更加合理。對于斗容量為0.25 m3的小型液壓挖掘機，單只動臂液壓缸即可滿足工作要求。具體結(jié)構(gòu)如圖2.2所示。（四）鏟斗與鏟斗油缸的連接方式本方案中采用六連桿的布置方式，相比四連桿布置方式而言在相同的鏟斗油缸行程下能得到較大的鏟斗轉(zhuǎn)角，改善了機構(gòu)的傳動特性。該布置中1桿與2桿的鉸接位置雖然使鏟斗的轉(zhuǎn)角減少但保證能得到足夠大的鏟斗平均挖掘力。如圖2.4所示。2331-斗桿； 2-連桿機構(gòu)； 3-鏟斗圖2.4 鏟斗連接布置示意圖（五）鏟斗的結(jié)構(gòu)選擇鏟斗結(jié)構(gòu)形狀和參數(shù)的合理選擇對挖掘機的作業(yè)效果影響很大，合適的鏟斗應滿足以下要求：1、有利于物料的自由流動。鏟斗內(nèi)壁不宜設置橫向凸緣、棱角等。斗底的縱向剖面形狀要適合于各種物料的運動規(guī)律。2、要使物料易于卸盡。3、為使裝進鏟斗的物料不易于卸出，鏟斗的寬度與物料的粒徑之比應大于4，大于50時，顆粒尺寸不考慮，視物料為均質(zhì)。綜上考慮，選用小型挖掘機常用的鏟斗結(jié)構(gòu)，基本結(jié)構(gòu)如圖2.5所示。圖2.5 鏟斗斗齒的安裝連接采用橡膠卡銷式，結(jié)構(gòu)示意圖如2.6所示。1-卡銷 ；2 橡膠卡銷；3 齒座； 4斗齒圖2.6 卡銷式斗齒結(jié)構(gòu)示意圖(六) 原始幾何參數(shù)的確定1、動臂與斗桿的長度比K1由于所設計的挖掘機適用性較強，作業(yè)對象明確，一般不替換工作裝置，故取中間比例方案，K1取在1.52.0之間?？紤]到K1值大，工作裝置結(jié)構(gòu)重心離機體近。初步選取K1=2，即l1 / l2=2。2、鏟斗斗容與主參數(shù)的選擇斗容量在任務書中已經(jīng)給出：q =0.25 m3按經(jīng)驗公式和比擬法初選：l3=900mm，鏟斗平均寬度B=800mm，鏟斗切削半徑R= l3=900mm,鏟斗裝滿轉(zhuǎn)角。3、工作裝置液壓系統(tǒng)主參數(shù)的初步選擇各工作油缸的缸徑選擇要考慮到液壓系統(tǒng)的工作壓力和“三化“要求。初選動臂油缸內(nèi)徑D1=125mm，活塞桿的直徑d1=80mm。斗桿油缸的內(nèi)徑D2=90mm，活塞桿的直徑d2=63mm。鏟斗油缸的內(nèi)徑D3=100mm，活塞桿的直徑d3=70mm。按經(jīng)驗公式初選各油缸全伸長度與全縮長度之比：1=2=3=1.6。參照任務書的要求選擇工作裝置液壓系統(tǒng)的工作壓力P=20MPa，閉鎖壓力Pg=21MPa。三、工作裝置運動學分析（一） 動臂運動分析動臂油缸的最短長度；動臂油缸的伸出的最大長度；A：動臂油缸的下鉸點；B：動臂油缸的上鉸點；C：動臂的下鉸點.圖3.1 動臂擺角范圍計算簡圖動臂擺角1是L1的函數(shù)。動臂上任意一點在任一時刻的坐標值也都是L1的函數(shù)。如圖3.1所示，圖中動臂油缸的最短長度；動臂油缸的伸出的最大長度；動臂油缸兩鉸點分別與動臂下鉸點連線夾角的最小值；動臂油缸兩鉸點分別與動臂下鉸點連線夾角的最大值；A：動臂油缸的下鉸點；B：動臂油缸的上鉸點；C：動臂的下鉸點。則有：在三角形ABC中： (3-1)圖3.2 F、C點坐標計算簡圖在三角形BCF中： (3-2)由圖3.2所示的幾何關系，可得到21的表達式： (3-3)當F點在水平線CU之下時21為負，否則為正。F點的坐標為 XF = l30+l1cos21 YF = l30+l1sin21 (3-4)C點的坐標為 YC = YA+l5sin11 (3-5)動臂油缸的力臂e1 (3-6)顯然動臂油缸的最大作用力臂e1max= l5(二）斗桿的運動分析如下圖3.3所示，D點為斗桿油缸與動臂的鉸點點，F(xiàn)點為動臂與斗桿的鉸點，E點為斗桿油缸與斗桿的鉸點。斗桿的位置參數(shù)是l2，這里只討論斗桿相對于動臂的運動，即只考慮L2的影響。D-斗桿油缸與動臂的鉸點點； F-動臂與斗桿的鉸點；E-斗桿油缸與斗桿的鉸點； 2-斗桿擺角.圖3.3 斗桿機構(gòu)擺角計算簡圖在三角形DEF中 (3-7)由上圖的幾何關系知斗桿相對于動臂的擺角范圍2max2max =2 max-2min (3-8)則斗桿的作用力臂 (3-9)顯然斗桿的最大作用力臂e2max = l9，此時。（三）鏟斗的運動分析鏟斗相對于XOY坐標系的運動是L1、L2、L3的函數(shù)，現(xiàn)討論鏟斗相對于斗桿的運動，如圖3-4所示，G點為鏟斗油缸與斗桿的鉸點，F(xiàn)點為斗桿與動臂的鉸點Q點為鏟斗與斗桿的鉸點，v點為鏟斗的斗齒尖點，K點為連桿與鏟斗的餃點，N點為曲柄與斗桿的鉸點，M點為鏟斗油缸與曲柄的鉸點，H點為曲柄與連桿的鉸點。圖3.4 鏟斗連桿機構(gòu)傳動比計算簡圖1、鏟斗連桿機構(gòu)傳動比i利用圖3.4，可以求得以下參數(shù)：在三角形HGN中32 = GMN = - MNG - MGN = -22-30 (3-10)在三角形HNQ中 (3-11)在三角形QHK中 (3-12)在四邊形KHNQ中NHK=NHQ+QHK (3-13)鏟斗油缸對N點的作用力臂r1 (3-14)連桿HK對N點的作用力臂r2r2 = l13Sin NHK 連桿HK對Q點的作用力臂r3 (3-15)連桿機構(gòu)的總傳動比i (3-16)顯然3-17式中可知，i是鏟斗油缸長度L3的函數(shù)，用L3min代入可得初傳動比i0，L3max代入可得終傳動比iz。2、鏟斗相對于斗桿的擺角3鏟斗的瞬時位置轉(zhuǎn)角為 (3-17)其中，在三角形NFQ中 (3-18)當鏟斗油缸長度L3分別取L3max和L3min時，可分別求得鏟斗的最大和最小轉(zhuǎn)角3max和3min，于是得鏟斗的擺角范圍: 3 = 3max-3min (3-19)3、斗齒尖運動分析見圖3.5所示，斗齒尖V點的坐標值XV和YV，是L1 、L2、L3的函數(shù)只要推導出XV和YV的函數(shù)表達式，那么整機作業(yè)范圍就可以確定，現(xiàn)推導如下:由F點知：32= CFQ= 2 3 4 6 2 (3-20)在三角形CDF中：DCF由后面的設計確定，在DCF確定后則有： (3-21) (3-22) (3-23)在三角形DEF中 圖3.5 齒尖坐標方程推導簡圖1則可以得斗桿瞬間轉(zhuǎn)角2 (3-24)4、6在設計畫圖中確定。由三角形CFN知：l28 = Sqr（l162 + l12 - 2cos32l16l1） (3-25)由三角形CFQ知：l23 = Sqr（l22 + l12 - 2cos32l2l1） (3-26)由Q點知：35= CQV= 2 33 24 10 (3-27)在三角形CFQ中： (3-28)在三角形NHQ中： (3-29)在三角形HKQ中： (3-30)在四邊形HNQK：NQH =24 + 26 (3-31)20 = KQV，其在后面的設計中確定。在列出以上的各線段的長度和角度之間的關系后，利用矢量坐標我們就可以得到各坐標點的值。（四） 特殊工作位置計算1、最大挖掘深度H1maxNH-搖臂；HK-連桿；C-動臂下鉸點；A -動臂油缸下鉸點；B-動臂與動臂油缸鉸點；F-動臂上鉸點；D-斗桿油缸上鉸點；E-斗桿下鉸點；G-鏟斗油缸下鉸點；Q-鏟斗下鉸點；K-鏟斗上鉸點；V-鏟斗斗齒尖.圖3.6 最大挖掘深度計算簡圖如圖3.6示，當動臂全縮時，F(xiàn), Q, V三點共線且處于垂直位置時，得最大挖掘深度為: H1max = YV = YFmin l2 l3 = YC + L1 Sin2 1min l2 l3 = YC + l1 Sin（1 20 11） l2 l3 (3-32)2、最大卸載高度H3maxNH-搖臂；HK-連桿；C-動臂下鉸點；A -動臂油缸下鉸點；B-動臂與動臂油缸鉸點；F-動臂上鉸點；D-斗桿油缸上鉸點；E-斗桿下鉸點；G-鏟斗油缸下鉸點；Q-鏟斗下鉸點；K-鏟斗上鉸點；V-鏟斗斗齒尖圖3.7 最大卸載高度計算簡圖如圖3.7所示，當斗桿油缸全縮，動臂油缸全伸時，QV連線處于垂直狀態(tài)時，得最大卸載高度為: (3-33)3、水平面最大挖掘半徑R1maxNH-搖臂；HK-連桿；C-動臂下鉸點；A -動臂油缸下鉸點；B-動臂與動臂油缸鉸點；F-動臂上鉸點；D-斗桿油缸上鉸點；E-斗桿下鉸點；G-鏟斗油缸下鉸點；Q-鏟斗下鉸點；K-鏟斗上鉸點；V-鏟斗斗齒尖圖3.8 停機面最大挖掘半徑計算簡圖如圖3.8所示，當斗桿油缸全縮時，F(xiàn)、 Q、V三點共線，且斗齒尖v和鉸點C在同一水平線上，即YC = YV，得到最大挖掘半徑R1max為:R1max=XC+L40 (3-34)式中：L40 = Sqr（L1+L2+L3）2 2（L2+L3）L1COS32max (3-35)4、最大挖掘半徑R2max NH-搖臂；HK-連桿；C-動臂下鉸點；A -動臂油缸下鉸點；B-動臂與動臂油缸鉸點；F-動臂上鉸點；D-斗桿油缸上鉸點；E-斗桿下鉸點；G-鏟斗油缸下鉸點；Q-鏟斗下鉸點；K-鏟斗上鉸點；V-鏟斗斗齒尖圖5.1 最大挖掘半徑時工作裝置結(jié)構(gòu)簡圖最大挖掘半徑時的工況是水平面最大挖掘半徑工況下C、V連線繞C點轉(zhuǎn)到水平面而成的。通過兩者的幾何關系，我們可計算得到：l 30 = 350mm ；l 40 = 5650mm。5、最大挖掘高度H2max最大挖掘高度工況是最大卸載高度工況中鏟斗繞Q點旋轉(zhuǎn)直到鏟斗油缸全縮而形成的。具體分析方法和最大卸載高度工況的分析類似。四、 挖掘阻力分析（一）轉(zhuǎn)斗挖掘阻力計算挖掘阻力可分為切向分力與法向分力,其中法向分力相對很小，一般為 (4-1) (4-2)在式(4-2)中，F(xiàn)1 切削阻力的切向分力；C土壤的硬度系數(shù)，對不同的土壤條件取值不同，這里設挖機用于級土壤的挖掘，取值為90；R鏟斗與斗桿鉸點到斗齒尖距離，即轉(zhuǎn)斗切削半徑其在前面已經(jīng)初步確定，取值為90 cm；max挖掘過程中鏟斗總轉(zhuǎn)角的一半；現(xiàn)初定總轉(zhuǎn)角為110，則max = 55某一挖掘位置處轉(zhuǎn)斗的瞬時轉(zhuǎn)角，B切削刃寬度影響系數(shù)，B = 1 + 2.6b = 1 + 2.60.8 = 3.08；A切削角變化影響系數(shù)，取A = 1.3.；Z帶有斗齒的系數(shù)，取Z =0.75；X斗側(cè)壁厚影響系數(shù)，X = 1+0.03S,其中S為側(cè)壁厚度，單位為cm 。初步設計時取X = 1.15 ；D切削刃擠壓土壤的力，根據(jù)經(jīng)驗統(tǒng)計和斗容量的大小選取D = 0.8 104N。當時，出現(xiàn)轉(zhuǎn)斗挖掘最大切向分力，其值為： (4-3)將各參數(shù)代入式(4-3)得 轉(zhuǎn)斗平均挖掘阻力按平均挖掘深度下的阻力計算，平均切削厚度為 (4-4)平均挖掘阻力為 (4-5) 將各參數(shù)代入上式得（二）斗桿挖掘阻力計算斗桿在挖掘過程中總轉(zhuǎn)角一般為，現(xiàn)取。斗齒尖的行程實際上是斗桿轉(zhuǎn)角所對應的弧長，根據(jù)經(jīng)驗公式有 （4-6）斗桿挖掘時切削半徑，斗桿與動臂鉸點至斗齒尖距離，單位m斗桿挖掘時切削厚度按如下公式計算 （4-7）q鏟斗容量，B鏟斗切削寬度m斗桿挖掘阻力計算公式如下： （4-8）式（4-8）中為挖掘阻力比，由附表010查得，對于級土取,對于，初步設計時取,將各參數(shù)代入式（4-8）得 取整為，斗桿挖掘阻力比轉(zhuǎn)斗挖掘阻力要小一些，這是由于斗桿挖掘行程較長，切削厚度較小的緣故。五、基本尺寸的確定（一）斗形參數(shù)的確定斗容量q ：在設計任務書中已給出q = 0.25 m3平均斗寬B：在設計任務書中已給出B = 0.8 m挖掘半徑R：按經(jīng)驗統(tǒng)計和參考同斗容的其它型號的機械，初選R = 900mm 轉(zhuǎn)斗挖掘裝滿轉(zhuǎn)角（2）：R、B及2三者與q之間有以幾何關系q = 0.5 R2B（2-Sin2）KS在上式中：KS為土壤的松散系數(shù)，近似取值為1.25。將q = 0.25 m3和B = 0.8m代入上式有：鏟斗兩個鉸點K、Q之間的間距l(xiāng)24和l3的比值k2的選?。簂24太大將影響機構(gòu)的傳動特性，太小則影響鏟斗的結(jié)構(gòu)剛度3，一般取特性參數(shù)。初選特性參數(shù)k2 = 0.3。一般取。由于鏟斗的轉(zhuǎn)角較大，而k2的取值較小，故初選。(二）動臂機構(gòu)參數(shù)的選擇1、1與A點坐標的選取初選動臂彎角。由經(jīng)驗統(tǒng)計和參考其它同斗容機型，初選特性參數(shù)k3 = 1.65（k3 = L42/L41）鉸點A坐標的選擇：由底盤和轉(zhuǎn)臺結(jié)構(gòu)，并結(jié)合同斗容其它機型的測繪，初選：XA = 560 mm ；YA = 700mm 2、 l1與l 2的選擇 經(jīng)統(tǒng)計分析，最大挖掘半徑R1值與l1+l2+l3的值很接近，由已給定的最大挖掘半徑R1、已初步選定的l3和k1，結(jié)合如下經(jīng)驗公式：； 式中： l1為動臂長， l 2為斗桿長，k1為動臂斗桿長度比將各參數(shù)代入上式得： ；3、 l41與l42的計算如圖5.1所示，在三角形CZF中： NH-搖臂；HK-連桿；C-動臂下鉸點；A -動臂油缸下鉸點；B-動臂與動臂油缸鉸點；F-動臂上鉸點；D-斗桿油缸上鉸點；E-斗桿下鉸點；G-鏟斗油缸下鉸點；Q-鏟斗下鉸點；K-鏟斗上鉸點；V-鏟斗斗齒尖圖5.1 最大挖掘半徑時工作裝置結(jié)構(gòu)簡圖l42 = k3l41 = 1.651407 = 2321 mm 4、l5的計算對于以反鏟為主的通用挖掘機要適當考慮其他的換用裝置（如正鏟、起重等），而且要求在地面以上作業(yè)時能有足夠的提升力矩，故初取k4 = 0.8511的取值對特性參數(shù)k4、最大挖掘深度H1max和最大挖高H2max均有影響，增大11會使k4減少或使H1max 增大，這符合反鏟作業(yè)的要求，初選。斗桿液壓油缸全縮時，CFQ =32 8最大，根據(jù)經(jīng)驗統(tǒng)計和便于計算，初選（32 8）max = 。由于采用單動臂液壓缸，因此BCZ的取值較大，初取BCZ = 如上圖5.1所示，在三角形CZF中：ZCF = 1 39 = - - = BCF = 2 =ZCF -ZCB由式（3-33）和式（3-34）有H3max = YC+ l1 Sin（1 20 11） l2 l3 (5-1) = YA+ l5 Sin11+ l1 Sin（1max 2 11）+ l2 Sin（1max+32 max 11 8 2 180） l3 H1max = l2 + l3 + l1 Sin（11 1min+ 2） l5 Sin11 YA ) (5-2)由式（5-1）、（5-2）有：H1max + H3max = l1 Sin（1max 2 11）+ l2 Sin（1max+ 32 max 11 8 2 180）+ l1 Sin（11 1min+ 2）+ l2 (5-3)令 A = 2+ 11 = + = B = A + （32 8）max = +（）=將A、B的值代入式（5-3）中有H1max + H3max l1 Sin（1max ） Sin（1min ） + l2 Sin （1max +）1=0 又由特性參數(shù) (5-4)則有 Sin1min = Sin1max 1 k4 = Sin1max1.36 (5-5) (5-6)將式（5-5）、式（5-6）代入到式（5-4）中得3500+3600-3400Sin（1max ） Sin（1min ）+l2Sin（1max +）1 = 0 解之： 1max = ； 1min = 由式(5-2)有H1max = l2 + l3 + l1 Sin（11- 1min +2）- l5 Sin11- YA l5 = l2 + l3 + l1 Sin（11- 1min + 2）- YA - H1max Sin11 = 1700 + 900 + 3400Sin（）- 800- 3500 Sin = 534.3mm1min與1max需要滿足以下條件 (5-7) (5-8)將1max 、1min 的值代入式（5-7）、式（5-8）中得： = 0.482 = 1.316而 (5-9) (5-10)、滿足5-9、5-10兩個經(jīng)驗條件，說明、的取值是可行的。 (5-11) (5-12) (5-13) 至此，動臂機構(gòu)的各主要基本參數(shù)已初步確定。（三） 動臂機構(gòu)基本參數(shù)的校核1、動臂機構(gòu)閉鎖力的校核由第四章的計算可知，轉(zhuǎn)斗的平均挖掘力由圖5-2知，最大挖掘深度時的挖掘阻力力矩M1J：M1J = （H1max + YC） (5-14)式中，YC為C點的Y軸坐標值將各參數(shù)代入式（5-14）得 M1J = 0.312 105（3.5+1.162）= 1.45105 N.m 動臂油缸的閉鎖力F1F1 = PgS1 （S1：動臂油缸小腔的作用面積） =2.1107（62.52 402）10 -6 = 1.5105 N 最大挖掘深度工作裝置自身重力所產(chǎn)生的力矩MG ：要求力矩，首先應該需要知道作用力和作用力臂。在此處，則是先要求出工作裝置各部分的重量，由經(jīng)驗統(tǒng)計，初步估計工作裝置的各部分重量如下：動臂G1 = 223kg 斗桿G2 = 179kg鏟斗G3 = 86kg 斗桿缸G4 = 55kg鏟斗缸G5 = 51kg 連桿機構(gòu)G6 = 17kg動臂缸G7 = 55kg 圖5.2 最大挖掘深度計算簡圖當處于最大挖掘深度時：1 = 1min = 由圖5.2有MG （G1/2 +G2 +G3 +G4 +G5 +G6+ G7）10 l1 cos (5-15) =（111.5+179 +86 +55 +51 +17+55）103.4 cos = 1.5104N.m 動臂油缸的閉鎖力與工作裝置重力所產(chǎn)生的力矩（對C點的矩）：M3 = F1l7 l5 Sin1min l1min + MG (5-16) = 21.51.459105 0.5343Sin40.51.109 + 1.5104 = 1.67105 N.m M1J = 1.45105 N.m 在式（5-16）中說明動臂油缸的閉鎖力與工作裝置重力所產(chǎn)生的力矩略大于平均挖掘阻力力矩，滿足工作要求。2、滿斗處于最大挖掘半徑時動臂油缸提升力矩的校核 NH-搖臂；HK-連桿；C-動臂下鉸點；A -動臂油缸下鉸點；B-動臂與動臂油缸鉸點；F-動臂上鉸點；D-斗桿油缸上鉸點；E-斗桿下鉸點；G-鏟斗油缸下鉸點；Q-鏟斗下鉸點；K-鏟斗上鉸點；V-鏟斗斗齒尖圖5.3 最大挖掘半徑時工作裝置結(jié)構(gòu)簡圖為方便計算，現(xiàn)將工作裝置劃分為二個部分，動臂、動臂液壓缸和斗桿液壓缸作為一部分，該部分重量以表示GB表示；其余的工作裝置構(gòu)件作為第二部分，重量以GG+D表示，于是有：GB=G1 +G4 +G7 =223 + 55 + 55 = 333kgGG+D =G2 +G3 +G5 +G6 = 179 + 86 + 51 + 17=333kg按經(jīng)驗公式取土的重量： GT = (1.6 1.8) q103 = 1.80.25 103 = 450kg當處于最大挖掘半徑時，工作裝置簡圖如圖5.3所示，則有：MZ = 9.8GB l1 /2 + GG+D（l1 + 0.7l2）+ GT （l1 + l2 + l3 /2） = 9.83333.42+ 333（3.4+0.71.7）+ 450（3.4+1.7-0.92） = 0.45105 N.m 動臂油缸的推力： F1 = P1 S1 = 210762.5210- 6 = 2.45105 N在如圖5.3所示，在三角形CAB中： (5-17)ACB =2 +11 +21 (5-18)將各參數(shù)分別代入式（5-17）和式（5-18）得 L1=1.542mL1 e1 = ACBCSinACB (5-19) 則此時動臂油缸提升力矩:MT = F1 e1= 2.451050.5054 =1.24105 N.m MZ = 0.45105 N.m 故鏟斗處于最大挖掘半徑時動臂油缸提升力矩滿足工作要求。3、滿斗處于最大高度時,動臂提升力矩的校核當斗桿在最大高度時的工況類似于圖3.7，此時動臂油缸全伸，斗桿油缸全縮。1 =1max = 32 =32max = 2 = 21 = 1-（2 + 11） 37 = 32 -（- 21）則工作裝置所受重力和土的重力所產(chǎn)生的載荷力矩MZ：MZ= (5-20)此時對于動臂油缸而言：L1 = L1max =1774 mm 1 =1max = 同式（5-19）的計算可求得此時的動臂油缸的力臂 此時動臂油缸的提升力矩MT可參考式（5-20）求得：MT = F1 e1 = 2010650210-60.388 = 0.61105 N.m MZ = 0.298105 N.m 說明滿斗處于最大高度時,動臂提升力矩滿足工作要求。E20（四）斗桿機構(gòu)基本參數(shù)的選擇E2ZD l92maxl8FD：斗桿油缸的下鉸點；E：鏟斗油缸的上鉸點；F動臂的上鉸點；2：斗桿的擺角；l9：斗桿油缸的最大作用力臂.圖5.4斗桿機構(gòu)基本參數(shù)計算簡圖取整個斗桿為研究對象，可得斗桿油缸最大作用力臂的表達式：e2max = l9 = F2d（l2 + l3 ）/ P2 = 2104 （1700+900）10 -3/2010645210-6 = 409 mm (5-21) 如圖5.4所示圖中，D：斗桿油缸的下鉸點；E：斗桿油缸的上鉸點；F動臂的上鉸點；2：斗桿的擺角；l9：斗桿油缸的最大作用力臂。斗桿油缸的初始位置力臂e20與最大力臂e2max有以下關系:e20 /e2max = l9 cos(2max /2)/l9 = cos (2max /2) (5-22)由5-22可知, 2max越大,則e20越小,即平均挖掘阻力越小.要得到較大的平均挖掘力,就要盡量減少2max,初取2max = 110由上圖5.43的幾何關系有:L2min = 2l9cSin (2max/2)/(2-1) = 2409Sin 55/(1.6 -1)= 1116.8 mm (5-23)L2max = L2min 2 = 1116.81.6= 1787 mm (5-24)l82 = L22min + l29 -2L2minl9cos( +2max)/2 = 1116.82+ 4092 + 21116.8409cos145 (5-25) l8 = 1470.6 mmEFQ取決于結(jié)構(gòu)因素和工作范圍,一般在130170之間，初定EFQ=160,動臂上DFZ也是結(jié)構(gòu)尺寸,按結(jié)構(gòu)因素分析,可初選DFZ=10。（五）鏟斗機構(gòu)基本參數(shù)的選擇1、轉(zhuǎn)角范圍由最大挖掘高度H2max和最大卸載高度H3max的分析，可以得到初始轉(zhuǎn)角：H2max-H3max = l3（Sin + 1.6） (5-26)將各參數(shù)代入式(5-26)得：5800-3600 = 900 （Sin + 1.6）， = 53最大轉(zhuǎn)角3max =V0QVZ，值太大會使斗齒平均挖掘力降低，常在150180之間選取，初選3max = 163。Kl292、鏟斗機構(gòu)其它基本參數(shù)的計算GL3Ml24l12FNQl21l2Vl3l12：搖臂的長度；l29：連桿的長度；l3：鏟斗的長度；l2：斗桿的長度；F：斗桿的下鉸點；G：鏟斗油缸的下鉸點；N：搖臂與斗桿的鉸接點；K：鏟斗的上鉸點；Q：鏟斗的下鉸點.圖5.5鏟斗機構(gòu)計算簡圖在圖5.5中有：l24 = KQ = k2 l3 = 0.3900 = 270mmL3max 與L3min 的確定：由第四章的計算可知轉(zhuǎn)斗平均挖掘阻力挖掘阻力F1P所做的W1p (5-27) 由圖5-5，鏟斗油缸推力所做的功W3：W3 = F3 （3-1）L3min = 2010650210-60.6L3min (5-28)由功的守恒知鏟斗油缸推力所做的功W3 應該等于鏟斗挖掘阻力所做的功W1p：即W3 = W1p (5-29)將5-27、5-28式代入5-29中計算可得：L3min = 849mm 圓整為850mm則L3max =3 L3min =1360mm剩余未選定的基本尺寸大部分為連桿機構(gòu)尺寸，其應滿足以下幾個條件：1）挖掘力的要求：鏟斗油缸的挖掘力應與轉(zhuǎn)斗最大挖掘阻力相適應，當斗齒尖處于V1時，斗桿油缸的理論挖掘力應不低于最大挖掘阻力的80% 。 即PD080% PD0max；當處于最大理論挖掘力位置時V1QV應為30。2）幾何相容。必須保證鏟斗六連桿機構(gòu)在l3全行程中任一瞬時都不會被破壞，即保證GFN、GHN以及四邊形HNQK在任何瞬時皆成立。3）l3全行程中機構(gòu)都不應出現(xiàn)死點，且傳動角應當在允許的范圍內(nèi)。根據(jù)以上三個方面的要求，通過經(jīng)驗公式和同斗容的其它機型的測繪對照，初步選定剩余的基本尺寸如下：HK = 352mm； HN = 407mm；NQ = 300mm； FN = l2-NQ = 1400mm； GF =432mm；預選GFN = 60則 GN 2 = FN 2 + GF 2 2COSGFNFNGF GN = 1242mm至此，工作裝置的基本尺寸均已初步確定。六、 工作裝置結(jié)構(gòu)設計整個工作裝置由動臂、斗桿、鏟斗及油缸和連桿機構(gòu)組成，要確定這些構(gòu)件的結(jié)構(gòu)尺寸，必須要對其結(jié)構(gòu)進行受力分析。要進行受力分析，首先要確定構(gòu)件最不利的工況，并找到在該工況下的危險截面，以作為受力分析的依據(jù)。但構(gòu)件在不利的工況下危險截面往往不止一個，這就需要分別計算出各危險截面尺寸再綜合考慮，取其中的最大值作為最終的尺寸。（一）斗桿的結(jié)構(gòu)設計1、斗桿的受力分析斗桿主要受到彎矩的作用，因此要找出斗桿中的最大彎矩進行設計計算。根據(jù)受力分析和以往的實驗表明，在鏟斗進行挖掘時，產(chǎn)生最大彎矩的工況滿足以下條件：1)動臂處于最低位置。即動臂油缸全縮。2)斗桿油缸的力臂最大。3)鏟斗齒尖在動臂與斗桿鉸點和斗桿與鏟斗鉸點的連線上。4)側(cè)齒挖掘時受到側(cè)向力Wk的作用。在這個工況下斗桿會存在最大彎矩，受到的應力也會最大。該工況的具體簡圖如圖6.1所示。取工作裝置為研究對象，如圖6.2所示。在該工況下存在的力有：工作裝置各部件所受到的重力Gi；作用在鏟斗上的挖掘阻力，包括切向阻力W1、法向阻力W2、側(cè)向阻力W3。VNH-搖臂；HK-連桿；C-動臂下鉸點；A -動臂油缸下鉸點；B-動臂與動臂油缸鉸點；F-動臂上鉸點；D-斗桿油缸上鉸點；E-斗桿下鉸點；G-鏟斗油缸下鉸點；Q-鏟斗下鉸點；K-鏟斗上鉸點；V-鏟斗斗齒尖圖6.1 斗桿危險工況時的工作裝置簡圖FNQPdW1HKW2G3HK-連桿 HN-搖臂N-搖臂與斗桿的鉸接點 Q-斗桿與鏟斗的鉸接點圖6.2 鏟斗受力分析簡圖當動臂油缸全縮時，通過前面的章節(jié)可以得出21 = 45，由圖6.1可知CF的向量可以表示為：FC = 3400COS（180-45）+Sin（180-45） = 3400（COS135+Sin135）由前面的章節(jié)計算結(jié)果知：ZFC =27，并初選DF = 1470mm。在DEF中DEF = 90COSEFD = EF/DF = 409/1470Design of machine and machine elementsMachine designMachine design is the art of planning or devising new or improved machines to accomplish specific purposes. In general, a machine will consist of a combination of several different mechanical elements properly designed and arranged to work together, as a whole. During the initial planning of a machine, fundamental decisions must be made concerning loading, type of kinematic elements to be used, and correct utilization of the properties of engineering materials. Economic considerations are usually of prime importance when the design of new machinery is undertaken. In general, the lowest over-all costs are designed. Consideration should be given not only to the cost of design, manufacture the necessary safety features and be of pleasing external appearance. The objective is to produce a machine which is not only sufficiently rugged to function properly for a reasonable life, but is at the same time cheap enough to be economically feasible. The engineer in charge of the design of a machine should not only have adequate technical training, but must be a man of sound judgment and wide experience, qualities which are usually acquired only after considerable time has been spent in actual professional work.Design of machine elements The principles of design are, of course, universal. The same theory or equations may be applied to a very small part, as in an instrument, or, to a larger but similar part used in a piece of heavy equipment. In no ease, however, should mathematical calculations be looked upon as absolute and final. They are all subject to the accuracy of the various assumptions, which must necessarily be made in engineering work. Sometimes only a portion of the total number of parts in a machine are designed on the basis of analytic calculations. The form and size of the remaining parts are designed on the basis of analytic calculations. On the other hand, if the machine is very expensive, or if weight is a factor, as in airplanes, design computations may then be made for almost all the parts. The purpose of the design calculations is, of course, to attempt to predict the stress or deformation in the part in order that it may sagely carry the loads, which will be imposed on it, and that it may last for the expected life of the machine. All calculations are, of course, dependent on the physical properties of the construction materials as determined by laboratory tests. A rational method of design attempts to take the results of relatively simple and fundamental tests such as tension, compression, torsion, and fatigue and apply them to all the complicated and involved situations encountered in present-day machinery. In addition, it has been amply proved that such details as surface condition, fillets, notches, manufacturing tolerances, and heat treatment have a market effect on the strength and useful life of a machine part. The design and drafting departments must specify completely all such particulars, must specify completely all such particulars, and thus exercise the necessary close control over the finished product. As mentioned above, machine design is a vast field of engineering technology. As such, it begins with the conception of an idea and follows through the various phases of design analysis, manufacturing, marketing and consumerism. The following is a list of the major areas of consideration in the general field of machine design: Initial design conception; Strength analysis; Materials selection; Appearance; Manufacturing; Safety; Environment effects; Reliability and life; Strength is a measure of the ability to resist, without fails, forces which cause stresses and strains. The forces may be; Gradually applied; Suddenly applied; Applied under impact; Applied with continuous direction reversals; Applied at low or elevated temperatures. If a critical part of a machine fails, the whole machine must be shut down until a repair is made. Thus, when designing a new machine, it is extremely important that critical parts be made strong enough to prevent failure. The designer should determine as precisely as possible the nature, magnitude, direction and point of application of all forces. Machine design is mot, however, an exact science and it is, therefore, rarely possible to determine exactly all the applied forces. In addition, different samples of a specified material will exhibit somewhat different abilities to resist loads, temperatures and other environment conditions. In spite of this, design calculations based on appropriate assumptions are invaluable in the proper design of machine. Moreover, it is absolutely essential that a design engineer knows how and why parts fail so that reliable machines which require minimum maintenance can be designed. Sometimes, a failure can be serious, such as when a tire blows out on an automobile traveling at high speeds. On the other hand, a failure may be no more than a nuisance. An example is the loosening of the radiator hose in the automobile cooling system. The consequence of this latter failure is usually the loss of some radiator coolant, a condition which is readily detected and corrected. The type of load a part absorbs is just as significant as the magnitude. Generally speaking, dynamic loads with direction reversals cause greater difficulties than static loads and, therefore, fatigue strength must be considered. Another concern is whether the material is ductile or brittle. For example, brittle materials are considered to be unacceptable where fatigue is involved. In general, the design engineer must consider all possible modes of failure, which include the following: Stress; Deformation; Wear; Corrosion; Vibration; Environmental damage; Loosening of fastening devices. The part sizes and shapes selected must also take into account many dimensional factors which produce external load effects such as geometric discontinuities, residual stresses due to forming of desired contours, and the application of interference fit joint. Selected from” design of machine elements”, 6th edition, m. f. sports, prentice-hall, inc., 1985 and “machine design”, Anthony Esposito, charles e., Merrill publishing company, 1975.Mechanical properties of materials The material properties can be classified into three major headings: (1) physical, (2) chemical, (3) mechanicalPhysical properties Density or specific gravity, moisture content, etc., can be classified under this category. Chemical propertiesMany chemical properties come under this category. These include acidity or alkalinity, react6ivity and corrosion. The most important of these is corrosion which can be explained in laymans terms as the resistance of the material to decay while in continuous use in a particular atmosphere. Mechanical properties Mechanical properties include in the strength properties like tensile, compression, shear, torsion, impact, fatigue and creep. The tensile strength of a material is obtained by dividing the maximum load, which the specimen bears by the area of cross-section of the specimen. This is a curve plotted between the stress along the This is a curve plotted between the stress along the Y-axis(ordinate) and the strain along the X-axis (abscissa) in a tensile test. A material tends to change or changes its dimensions when it is loaded, depending upon the magnitude of the load. When the load is removed it can be seen that the deformation disappears. For many materials this occurs op to a certain value of the stress called the elastic limit Ap. This is depicted by the straight line relationship and a small deviation thereafter, in the stress-strain curve (fig.3.1). Within the elastic range, the limiting value of the stress up to which the stress and strain are proportional, is called the limit of proportionality Ap. In this region, the metal obeys hookess law, which states that the stress is proportional to strain in the elastic range of loading, (the material completely regains its original dimensions after the load is removed). In the actual plotting of the curve, the proportionality limit is obtained at a slightly lower value of the load than the elastic limit. This may be attributed to the time-lagin the regaining of the original dimensions of the material. This effect is very frequently noticed in some non-ferrous metals. Which iron and nickel exhibit clear ranges of elasticity, copper, zinc, tin, are found to be imperfectly elastic even at relatively low values low values of stresses. Actually the elastic limit is distinguishable from the proportionality limit more clearly depending upon the sensitivity of the measuring instrument. When the load is increased beyond the elastic limit, plastic deformation starts. Simultaneously the specimen gets work-hardened. A point is reached when the deformation starts to occur more rapidly than the increasing load. This point is called they yield point Q. the metal which was resisting the load till then, starts to deform somewhat rapidly, i. e., yield. The yield stress is called yield limit Ay. The elongation of the specimen continues from Q to S and then to T. The stress-strain relation in this plastic flow period is indicated by the portion QRST of the curve. At the specimen breaks, and this load is called the breaking load. The value of the maximum load S divided by the original cross-sectional area of the specimen is referred to as the ultimate tensile strength of the metal or simply the tensile strength Au. Logically speaking, once the elastic limit is exceeded, the metal should start to yield, and finally break, without any increase in the value of stress. But the curve records an increased stress even after the elastic limit is exceeded. Two reasons can be given for this behavior: The strain hardening of the material; The diminishing cross-sectional area of the specimen, suffered on account of the plastic deformation. The more plastic deformation the metal undergoes, the harder it becomes, due to work-hardening. The more the metal gets elongated the more its diameter (and hence, cross-sectional area) is decreased. This continues until the point S is reached. After S, the rate at which the reduction in area takes place, exceeds the rate at which the stress increases. Strain becomes so high that the reduction in area begins to produce a localized effect at some point. This is called necking. Reduction in cross-sectional area takes place very rapidly; so rapidly that the load value actually drops. This is indicated by ST. failure occurs at this point T. Then percentage elongation A and reduction in reduction in area W indicate the ductility or plasticity of the material: A=(L-L0)/L0*100% W=(A0-A)/A0*100% Where L0 and L are the original and the final length of the specimen; A0 and A are the original and the final cross-section area. Selected from “testing of metallic materials”Quality assurance and control Product quality is of paramount importance in manufacturing. If quality is allowed deteriorate, then a manufacturer will soon find sales dropping off followed by a possible business failure. Customers expect quality in the products they buy, and if a manufacturer expects to establish and maintain a name in the business, quality control and assurance functions must be established and maintained before, throughout, and after the production process. Generally speaking, quality assurance encompasses all activities aimed at maintaining quality, including quality control. Quality assurance can be divided into three major areas. These include the following:Source and receiving inspection before manufacturing;In-process quality control during manufacturing;Quality assurance after manufacturing. Quality control after manufacture includes warranties and product service extended to the users of the product.Source and receiving inspection before manufacturing Quality assurance often begins ling before any actual manufacturing takes place. This may be done through source inspections conducted at the plants that supply materials, discrete parts, or subassemblies to manufacturer. The manufacturers source inspector travels to the supplier factory and inspects raw material or premanufactured parts and assemblies. Source inspections present an opportunity for the manufacturer to sort out and reject raw materials or parts before they are shipped to the manufacturers production facility. The responsibility of the source inspector is to check materials and parts against design specifications and to reject the item if specifications are not met. Source inspections may include many of the same inspections that will be used during production. Included in these are:Visual inspection;Metallurgical testing;Dimensional inspection;Destructive and nondestructive inspection;Performance inspection.Visual inspections Visual inspections examine a product or material for such specifications as color, texture, surface finish, or overall appearance of an assembly to determine if there are any obvious deletions of major parts or hardware.Metallurgical testing Metallurgical testing is often an important part of source inspection, especially if the primary raw material for manufacturing is stock metal such as bar stock or structural materials. Metals testing can involve all the major types of inspections including visual, chemical, spectrographic, and mechanical, which include hardness, tensile, shear, compression, and spectr5ographic analysis for alloy content. Metallurgical testing can be either destructive or nondestructive.Dimensional inspection Few areas of quality control are as important in manufactured products as dimensional requirements. Dimensions are as important in source inspection as they are in the manufacturing process. This is especially critical if the source supplies parts for an assembly. Dimensions are inspected at the source factory using standard measuring tools plus special fit, form, and function gages that may required. Meeting dimensional specifications is critical to interchangeability of manufactured parts and to the successful assembly of many parts into complex assemblies such as autos, ships, aircraft, and other multipart products.Destructive and nondestructive inspection In some cases it may be necessary for the source inspections to call for destructive or nondestructive tests on raw materials or p0arts and assemblies. This is particularly true when large amounts of stock raw materials are involved. For example it may be necessary to inspect castings for flaws by radiographic, magnetic particle, or dye penetrant techniques before they are shipped to the manufacturer for final machining. Specifications calling for burn-in time for electronics or endurance run tests for mechanical components are further examples of nondestructive tests. It is sometimes necessary to test material and parts to destruction, but because of the costs and time involved destructive testing is avoided whenever possible. Examples include pressure tests to determine if safety factors are adequate in the design. Destructive tests are probably more frequent in the testing of prototype designs than in routine inspection of raw material or parts. Once design specifications are known to be met in regard to the strength of materials, it is often not necessary to test further parts to destruction unless they are genuinely suspect.Performance inspection Performance inspections involve checking the function of assemblies, especially those of complex mechanical systems, prior to installation in other products. Examples include electronic equipment subcomponents, aircraft and auto engines, pumps, valves, and other mechanical systems requiring performance evaluation prior to their shipment and final installation. Selected form “modern materials and manufacturing process”Electro-hydraulic drum brakesApplication The YWW series electro-hydraulic brake is a normally closed brake, suitable for horizontal mounting. It is mainly used in portal cranes, bucket stacker/reclaimersslewing mechanism.The YKW series electro-hydraulic brake is a normally opened brake, suitable for horizontal mounting, employing a thruster as actuator. with the foot controlling switch the operator can release or close the brake. It is mainly used for deceleration braking of portal cranesslewing mechanism. In a non-operating state the machinery can be braked by a manual close device.The RKW series brake is a normally opened brake, which is operated by foot driven hydraulic pump, suitable for horizontal mounting. Mainly used in the slewing mechanism of middle and small portal cranes. When needed, the brake is activated by a manual closed device. Main design featuresInterlocking shoes balancing devices (patented technology) constantly equalizes the clearance of brake shoes on both sides and made adjustment unnecessary, thus avoiding one side of the brake lining sticking to the brake wheel. The brake is equipped with a shoed autoaligning device.Main hinge points are equipped with self-lubricating bearing, making high efficiency of transmission, long service life. Lubricating is unnecessary during operation.Adjustable bracket ensure the brake works well.The brake spring is arranged inside a square tube and a surveyors rod is placed on one side. It is easy to read braking torque value and avoid measuring and computing.Brake lining is of card whole-piece shaping structure, easy to replace. Brake linings of various materials such as half-metal (non-asbestos) hard and half-hard, soft (including asbestos) substance are available for customers to choose.All adopt the companys new types of thruster as corollary equipment which work accurately and have long life. Hydraulic Power TransmissionThe Two Types Of Power Transmission In hydraulic power transmission the apparatus (pump) used for conversion of the mechanical (or electrical,thermal) energy to hydraulic energy is arranged on the input of the kinematic chain ,and the apparatus (motor) used for conversion of the hydraulic energy to mechanical energy is arranged on the output (fig.2-1) The theoretical design of the energy converters depends on the component of the bernouilli equation to be used for hydraulic power transmission. In systerms where, mainly, hydrostatic pressure is utilized, displacement (hydrostatic) pumps and motors are used, while in those where the hydrodynamic pressure is utilized is utilized gor power transmission hydrodynamic energy converters (e.g. centrifugal pumps) are used. The specific characteristic of the energy converters is the weight required for transmission of unit power. It can be demonstrated that the use of hydrostatic energy converters for the low and medium powers, and of hydrodynamic energy converters of high power are more favorite (fig.2-2). This is the main reason why hydrostatic energy converters are used in industrial apparatus. transformation of the energy in hydraulic transmission. 1. driving motor (electric, diesel engine);2. mechanical energy;3. pump; 4. hydraulic energy; 5. hydraulic motor; 6. mechanical energy; 7. load variation of the mass per unit power in hydrostatic and hydrodynamic energy converters 1、hydrostatic; 2.hydrodynamicOnly displacement energy converters are dealt with in the following. The elements performing converters provide one or several size. Expansion of the working chambers in a pump is produced by the external energy admitted, and in the motor by the hydraulic energy. Inflow of the fluid occurs during expansion of the working chamber, while the outflow (displacement) is realized during contraction. Such devices are